With increasing oil prices and greater dependency in America on imported oil, engines with improved fuel economy provide tremendous benefits. In addition, carefully controlling the type and quantity of emissions from an engine can be important.
Mounting concerns about global warming are pointing to excess emission of air pollutants from combustion of hydrocarbons. Controlled emission gases are presently carbon monoxide, and excess hydrocarbons, both caused by excessively rich combustion. Emission of carbon dioxide can also be substantially reduced by introducing other hydrocarbon fuels of a different hydrogen-carbon structure.
A substantial amount of fuel can be saved if the spark ignition (SI) engines can be made to operate on much leaner fuel-air ratios without a substantial loss in engine power and potential flame out. The thermal operating efficiencies of many engines are poor, and little progress has been made in improvements the last several years.
In Otto-cycle engines, fuel and air are mixed outside the combustion chamber and ignited by an electric spark after compression. This brings the local fuel-air mixture above the autoignition temperature to start the combustion, which then takes place over a small change in combustion chamber volume. In a Diesel-cycle air is compressed alone in the combustion chamber to a high pressure and temperature level. This brings the air temperature above the autoignition temperature, fuel is injected into the combustion chamber directly and atomized to penetrate part of the combustion volume. The fuel-air mixture is ignited by the hot air, and combustion takes place in the chamber during continued fuel injection and combustion chamber volume expansion, which simulates constant pressure combustion to some degree.
In an Otto-cycle engine, a relatively homogeneous fuel-air mixture penetrates the combustion chamber and is combusted almost completely according to the fuel-air mixture and the local mixture temperatures. In a Diesel-cycle engine, a stratified, locally rich, fuel-air mixture is enclosed by excess air, which receives heat from the compression of the air. It is therefore obvious that combustion in a Diesel engine can take place in an overall very much leaner fuel-air mixture than an Otto engine combustion chamber, where the combustion flame must penetrate the combustion chamber completely. The entire fuel-air mixture must be within the flammability limit and above the autoignition temperature to consume all the fuel. The fuel-air mixture is compressed together in an Otto engine. Care must therefore be taken to prevent a premature start to combustion, caused by hot spots or excessive compression temperature above the autoignition temperature level. This makes it almost impossible to use a conventional Otto engine cycle in adiabatic or near adiabatic type of operation, where the combustion chamber wall temperature spots may reach autoignition levels.
The problem of premature autoignition or pre-ignition in an Otto engine is solved by using high octane fuels for combustion. FIG. 1 from Technology Reference (Tech. Ref.) 1 shows autoignition temperatures for unsaturated mixtures of low octane JP-4 and high octane AVGAS 115/145 and air at atmospheric pressure versus low flow velocities. For saturated mixtures at stagnant or low flow velocities the autoignition temperatures are lower. The figure shows the autoignition temperatures of the high octane fuel-air mixture to be some 200 degrees Fahrenheit higher than the low octane one. These values are typical for groups of similar fuels. The figure also shows that the fuel-air mixture flow velocity can compensate for lack of octane rating. Ignition delays for the low octane fuel show about 10 seconds at the lowest temperature level without flow. This reduces to 0.2 second at 1200 degrees Fahrenheit at fuel-air mixture flow velocities of about 18 ft/sec. Combustion time at constant pressure combustion is normally 30 times longer than the ignition delay, which suggests a very slow reaction. The important message here is that the combustion rate is enhanced substantially when conducted in a flow.
FIGS. 2 and 3 in the illustrations from Tech. Ref. 2 show the engine thermal efficiency and indicated power in a single cylinder reciprocating piston engine in Otto-cycle operation as functions of equivalence ratios for methanol and gasoline fuels. FIGS. 2 and 3 show that a standard mixture of gasoline and air will not ignite and burn beyond an equivalence ratio of about 0.8 unless turbulence is introduced. In that case, the flammability range may improve to an equivalence ratio of about 0.7 by improved mixing and with turbulence. Methanol, however, in the standard mixture will ignite and burn to an equivalence ratio of about 0.68, and for an improved mixture with turbulence to an equivalence ratio of about 0.6. There are some differences between gasoline and methanol in combustion performance. According to FIGS. 2 and 3, the stoichiometric mixture in the shown engine is found at an air-fuel ratio of 14.5 by mass of gasoline, while methanol has a stoichiometric mixture of 6.5. The flammability range of gasoline is given as 0.6 to 3.8 in terms of equivalence ratio, and for methanol as 0.45 to 4.2. More important might be the laminar flame speed, which for gasoline is given as 0.37 ft/sec, and for methanol 0.52 ft/sec. The adiabatic flame temperatures are about the same, and the heats of combustion are in the same ratio as the stoichiometric fuel-air ratios.
FIG. 2 further shows that some improvement in thermal efficiency is available at lower equivalence ratio operations. This is at the expense of indicated power, as seen from FIG. 3.
FIGS. 2 and 3 of the illustrations show little improvement in the lean flammability limit in a single cylinder reciprocating piston internal combustion engine due to compression of the fuel-air mixture compared with standard values. The values of these figures compare with values cited for the same fuels at standard conditions in chemical handbooks as described in the Background section of this disclosure. Introduction of turbulence and flow into the fuel-air mixture on the other hand extended the low flammability limits to lower equivalence ratios. The level of turbulence available in a piston engine is very limited. If a high degree of turbulence is sought, this can only be achieved with a very high flow velocity. Such a high flow velocity can only be reached in a closed vessel combustion chamber when the combustion chamber moves at a substantial velocity relative to the combustion chamber boundaries. This type of movement was introduced to a very moderate degree in the Wankel engine, but this engine suffered from slow combustion probably due to low ignition temperature and positioning of the igniter plug.
In the Wankel engine, the fuel-air mixture moves at travel speeds up to 30 ft/sec relative to the stator. In the combustion chambers in a gas turbine engine the flow velocity is rarely more than 70 ft/sec.
Information and data used in this disclosure are based on data and illustrations taken from the cited Technology References (Tech. Refs.) to describe and substantiate the technology basis for the observations made and the methods used.